Axial Piston Compressor

ABSTRACT

Axial piston compressor, especially for motor vehicle air-conditioning systems, having a tilt plate ( 2 ), especially a ring-shaped tilt plate, which is variable in terms of its inclination with respect to a drive shaft ( 1 ) and which is driven in rotation by the drive shaft ( 1 ) and is in articulated connection with at least one supporting element ( 6 ) arranged at a spacing from the drive shaft ( 1 ) and rotating together therewith, the pistons in each case having an articulated arrangement with which the tilt plate ( 2 ) is in sliding engagement, and the supporting element ( 6 ) being arranged at the radially outer end of a force transmission element ( 7 ) which rotates together with the drive shaft ( 1 ) and is fixed in the latter so as to be non-displaceable in the radial direction, wherein the force transmission element ( 7 ) is mounted in the drive shaft ( 1 ) so that it can rotate about its longitudinal axis.

The present invention relates to an axial piston compressor, especially a compressor for motor vehicle air-conditioning systems, in accordance with the preamble of claim 1.

In the field of compressor drive mechanisms, a trend is beginning to emerge that, in the case of compressors having variable piston stroke, increasing use is being made of tilt plates in the form of a tilt ring, that is to say ring-shaped tilt plates, with a tilt-providing articulation necessary for tilting of the plate being substantially integrated into the ring-shaped tilt plate. For example, there is known, from EP 0 964 997 B1, a compressor in which the stroke movement of the pistons is accomplished by means of engagement—in an engagement chamber—of a ring plate oriented on a slant to the machine shaft. The engagement chamber is provided adjacent to the enclosed-hollow space of the piston. For sliding engagement that is substantially free from play in any slanting position of the tilt plate or tilt ring there are provided on both sides, between it and the spherically curved inner wall of the engagement chamber, spherical segments, so-called sliding blocks, so that the tilt ring slides between them as it revolves.

The drive is transmitted from the drive shaft to the tilt ring by means of a pin for conjoint movement which is attached to the drive shaft and the spherical head of which engages in a radial bore in the tilt ring, the position of the head of the member for conjoint movement being selected so that its centre-point coincides with that of the spherical segments. In addition, that centre-point is located on a circular line which connects the geometric axes of the seven pistons with one another and, moreover, on a circular line which connects the centre-points of the spherical articulation members of the pistons. By that means, the upper dead-centre position of the pistons is determined and a minimum clearance volume is ensured. The head shape of the free end of the member for conjoint movement makes it possible for the inclination of the tilt plate to change by means of the fact that the head of the member for conjoint movement forms a bearing body for a tilting movement of the tilt plate which changes the stroke distance of the pistons.

A further precondition for tilting of the tilt plate is the displaceability of its mounting axis in the direction of the drive shaft. For the purpose, the mounting axis is formed by two mounting pins mounted on the same axis on each side of a sliding sleeve, which mounting pins are additionally mounted in radial bores in the tilt plate. For the purpose, the sliding sleeve preferably has mounting sleeves on each side, which span the annular space between the sliding sleeve and the tilt plate in the manner of spokes.

The limitation on the displaceability of the mounting axis and, as a result, the maximum angled position of the tilt plate results from the pin for conjoint movement, by virtue of the fact that the latter passes through an elongate hole provided in the sliding sleeve so that the sliding sleeve meets end stops at the ends of the elongate hole. The force for the change in the angle of the tilt plate and, therefore, for regulation of the compressor results from the sum of the pressures acting against one another in each case on each side of the pistons, so that this force is dependent on the pressure in the drive mechanism chamber. In accordance with the prior art, the pressure in the drive mechanism chamber can be regulated between a high pressure and a low pressure and consequently affects the balance of forces at the tilt plate, which influences the inclination of the latter. The position of the sliding sleeve can moreover be influenced by springs which, in various variants, are likewise included in the prior art.

Furthermore, the position of the sliding sleeve, which position governs the delivery output, is also determined by the forces of inertia acting on the tilt plate; the position of the tilt plate, that is to say its angle of tilt or slant, changes with increasing speed of rotation. In the case of modern compressors, the trend is towards using tilt plates having moments of inertia such that they bring about a reduction in the stroke distance of the pistons and therefore a reduction in delivery output when the speed of rotation increases.

However, what is problematic in the arrangement explained hereinbefore is the high Hertzian stress in the region of the head of the member for conjoint movement and the tilt plate (system: sphere/cylinder) and the take-up of the (axial) reaction forces due to the gas force on the pistons and the forces due to the torque to be transmitted to the tilt plate.

A compressor similar to the compressor known from EP 0 964 997 B1 is known from JP 2003-269330 AA, although in that compressor a total of two members for conjoint movement are used.

It is important to the kinematics according to the two mentioned publications, that is to say to the kinematics in the case of the subject-matter of EP 0 964 997 B1 and JP 2003-269330 AA, that the head of the member for conjoint movement centrally coincides with the centre-point of the sliding blocks of the pistons and that the position of the centre-point of the head of the member for conjoint movement is at the same time approximately tangential to the reference circle of the central axes of the pistons.

Added to the afore-mentioned disadvantageous characteristics is the fact that the subject-matter of EP 0 964 997 B1 and of JP 2003-269330 AA has a very complicated structural arrangement, which results in a high number of parts and therefore costs, and in addition the mounting by means of two members for conjoint movement is over-determined and therefore susceptible to wear, and the strength of the components, especially due to the fact that a hole is introduced into the shaft, has to be regarded as rather low.

A further compressor is known from DE 101 52 097 A1, differing considerably from the subject-matter of the publications discussed hereinbefore. In the case of the subject-matter according to DE 101 52 097 A1, the member for conjoint movement, in particular the spherical head of the member for conjoint movement, is replaced by a hinge pin or spindle. This is, however, integrated into the tilt plate from the outside and fastened using a cup-shaped disc for conjoint movement which is a component of the drive shaft assembly. The subject-matter of DE 101 52 097 A1 also has a complicated structural arrangement; in addition it has to be borne in mind that a large imbalance can come about, depending on the angle of tilt. This promotes wear on the compressor and as a result reduces its service life.

A further compressor is known from FR 278 21 26 A1, which has a member for conjoint movement extending out from the drive shaft radially and engaging in the tilt plate. In similar manner to the solution according to DE 101 52 097 A1, the tilt plate in this arrangement is also fixed to the member for conjoint movement in radial extension. In this there also lies a central difference from the subject-matter of EP 0 964 997 B1 and JP 2003-269330 AA. Whereas in the latter cases the mounting point of the head of the member for conjoint movement in the tilt plate undergoes relative movement in the guideway (bore) in the tilt plate because the tilt plate performs the rotary movement in an articulation lying on the shaft axis, the rotary movement in the case of the arrangements according to FR 278 21 26 A1 and DE 101 52 097 A1 is accomplished in the lateral articulation of the tilt plate.

In the unpublished Patent Application DE 102 00 404 1645 belonging to the present Applicant, there is proposed a member for conjoint movement which is displaceably mounted in the shaft. As a result, the transmission of force between the head of the member for conjoint movement and the tilt plate can be accomplished optimally (force transmission as a result of area-wise contact). However, the displacement of the member for conjoint movement in the shaft can be problematic because high forces have to be taken up there owing to the bending moment and the parts therefore have to be of very rigid construction. This rigid construction causes the compressor to have an increased mass.

Finally, from DE 103 154 77 A1 there is known a compressor of the tilt plate/member for conjoint movement constructional type wherein the member for conjoint movement does not transmit any torque. This feature in addition also applies to preferred arrangements of DE 102 00 404 1645. The conjoint movement function is restricted to providing support for the piston forces acting axially on the tilt plate, the torque being delivered by further force transmission elements independent of the member for conjoint movement. As a result, the forces acting on the member for conjoint movement are lower because, as already mentioned, no torque is transmitted. The advantage of this approach lies in the fact that the forces or contact pressure due to the forces applied (because of the fact that these forces are relatively low) do not cause any excessive deformation at and in the member for conjoint movement, as a result of which the member for conjoint movement can be of correspondingly lightweight construction and tilting of the tilt plate can be accomplished in a relatively hysteresis-free manner. However, a disadvantageous effect can be that the spherical head of the member for conjoint movement is located in a relatively large recess in the tilt plate. As a result, the Hertzian stress can or must be described by a plane/sphere geometric pairing, which is relatively disadvantageous because it causes a high degree of Hertzian stress.

Starting from the prior art explained hereinbefore, the problem of the present invention is to provide a compressor in which the head of the member for conjoint movement or its supporting element can take up forces over as great an area as possible (low Hertzian stress), whilst at the same time avoiding over-determination of force transmission functions, that is to say restriction.

The problem is solved by a compressor having the features according to patent claim 1 and 3.

A compressor according to the invention has a tilt plate, especially a ring-shaped tilt plate, which is variable in terms of its inclination to a drive shaft and which is driven in rotation by the drive shaft and is in articulated connection with at least one supporting element arranged at a spacing from the drive shaft and rotating together therewith. The pistons of the axial piston compressor in each case have an articulated arrangement with which the tilt plate is in sliding engagement. The supporting element is arranged at the radially outer end of a force transmission element which rotates together with the drive shaft and which is fixed in the latter so as to be non-displaceable in the radial direction, a fundamental point of the invention being that the force transmission element is mounted in the drive shaft so that it can rotate about its longitudinal axis. By this means it is ensured that no undesirable moments, especially torsional moments, act on the force transmission element and also on the supporting element and result in increased wear.

In a preferred embodiment of a compressor according to the invention, the supporting element and force transmission element serve substantially only for providing axial support for the pistons or support for the gas force, whereas the torque transmission between the drive shaft and tilt ring is accomplished by a specific arrangement independent of the supporting element and force transmission element. This arrangement is especially an articulated connection between the drive shaft and the tilt plate. As a result of separation of the arrangement serving for providing axial support for the pistons or support for the gas force from an arrangement dedicated to driving the tilt plate or tilt ring, the arrangement serving for providing support for the gas force can firstly be slimmer and therefore of lighter construction, and furthermore there is obtained the advantage that the moments acting on the supporting element and the force transmission element can be reduced, which, as already mentioned hereinbefore, ensures less wear on the compressor according to the invention.

A further solution to the problem of the present invention is obtained when the supporting element in a compressor according to the preamble of patent claim 1 has a basic shape which in radial section is approximately rectangular, the “corners” being highly rounded especially with different radii, or also alternatively in the form of a compressed or deformed circle or also of an ellipse, which can in turn be deformed or compressed. As a result, the contact pressure or deformation in the region of the supporting element and tilt plate is advantageously influenced. Of course, a combination of the features of patent claims 1 and 3 is also possible

Those regions of the supporting element which are in contact with the tilt plate or tilt ring can be of at least partly cylindrical or barrel-shaped construction. As a result of a cylindrical or barrel-shaped contour, the tilt plate is approximately in line contact with the supporting element. It should be pointed out that the contour of the supporting element as described above can be machined to shape or ground by means of a shaping tool, which ensures simple and, as a result, economical manufacture.

In the case of a coolant compressor of the described mode of construction or in the case of a coolant compressor according to the preamble of patent claim 1, it is possible to differentiate between two moments in respect of the tilting of the tilt plate. These moments are, on the one hand, a tilting moment (in the tilting plane) and, on the other hand, a torsional moment, which acts perpendicular to the afore-mentioned tilting moment. The torsional moment arises inter alia because the maximum gas force at a piston arises at the moment of opening of the valve and not at the upper dead-centre of the piston. The resulting reaction force of all the pistons directs itself to a large extent to that piston which is in the state described. In a preferred embodiment, support for the torsional moment is provided in the region of the drive shaft, which is accomplished especially by means of an arrangement which is provided between a sliding sleeve mounted on the drive shaft in displaceable manner and the tilt plate. An arrangement of such a kind can consist of one or more cylindrical-pin-like element(s) or of supporting or contact surfaces. At this juncture it should be pointed out that the tilt plate is pivotally mounted on the afore-mentioned sliding sleeve, which is mounted so as to be axially displaceable along the drive shaft. As a result, as already mentioned hereinbefore, support is provided for a torsional moment acting in the region of the drive shaft. The force transmission element can furthermore have, at least over parts of its periphery, a shoulder in the region of the drive shaft and, additionally or alternatively, can comprise at its end remote from the supporting element a securing element, especially extending in the axial direction. The shoulder in the region of the drive shaft ensures that the force transmission element has a defined position in the drive shaft in a simple structural arrangement, and a securing element ensures secure holding in the drive shaft. The tilt plate is connected to the sliding sleeve and to the drive shaft preferably by means of drive pins, especially in a constructional form in which the force transmission element and the supporting element serve solely as support for the gas force. This ensures a reliable, structurally simple drive to the tilt plate, whereas at the same time the advantages of separating the drive from the provision of support against gas forces come into effect. For a secure hold, the drive pins can be introduced into the sliding sleeve or the tilt plate with a press fit. Furthermore, the drive pins can project into a recess, especially a groove, in the drive shaft, in which case furthermore a connecting element, especially a feather key, can be arranged between the drive shaft and the sliding sleeve, which connecting element allows transmission of forces and moments in a radial direction and is mounted in axially displaceable manner on the drive shaft. This ensures a relatively simple variant of the axial piston compressor according to the invention which can be produced using just a few individual parts.

A further preferred structurally simple embodiment of a compressor according to the invention is obtained when that end of the force transmission element which is remote from the supporting element projects through the drive shaft and into a longitudinal slot in the sliding sleeve in such a way that drive torque is transmitted from the drive shaft to the sliding sleeve by means of that end of the force transmission element which is remote from the supporting element.

In a particular embodiment, the supporting element is so constructed that it is located within a recess in the tilt plate in line contact with the latter, which ensures optimum Hertzian stress and also optimum force transmission. The height of the recess in the tilt plate can be equal to the sum of the radii of curvature of a radially outer and a radially inner contour, which ensures an ideal curve profile for the gas force support means. On that side of the tilt plate which is subject to greater loading by the gas force, the wall thickness in the region of the recess in the tilt plate is preferably greater than on that side which is subject to less loading, whilst furthermore at the same time the clearance volume is constant for all tilt angles of the tilt plate. Normally, that side of the tilt plate which is subject to greater loading by the gas force is the side facing the pistons. This structural measure increases the stability of the tilt plate, whilst at the same time the thinner wall thickness on that side which is subject to less loading makes it possible to save weight.

The drive shaft and the sliding sleeve can have flattened regions that correspond to one another, so that the sliding sleeve is mounted on the drive shaft in a manner ensuring conjoint rotation. This represents a simple structural measure for ensuring a reliable drive to the sliding sleeve. Furthermore, the tilt plate can have at least one flattened region which corresponds to a flattened region on the sliding sleeve, which ensures a reliable relative position of the two components with respect to one another.

The invention will be described hereinbelow with regard to further advantages and features by way of example and with reference to the accompanying drawings, in which:

FIG. 1 shows, in an exploded view, a tilt plate mechanism of a first preferred embodiment of a compressor according to the invention;

FIG. 2 shows, in longitudinal section, the embodiment according to FIG. 1 at a minimum tilt angle of the tilt plate;

FIGS. 3 a+b show the tilt plate mechanism according to FIG. 1 at a maximum tilt angle of the tilt plate—in longitudinal section (a) and in transverse section (b);

FIGS. 4 a+b show a detail of a gas force support means according to the invention in longitudinal section (a) and diagrammatic views of the gas force support means again in longitudinal section (b);

FIG. 5 shows, in longitudinal section, a second preferred embodiment of a gas force support means of a compressor according to the invention;

FIG. 6 shows, again in longitudinal section, a third preferred embodiment of a gas force support means of a compressor according to the invention; and

FIGS. 7 a+b show a tilt plate mechanism of the first preferred embodiment—in transverse section (a) and in longitudinal section (b).

All preferred embodiments of the compressor according to the invention comprise (not shown in the drawings) a housing, a cylinder block and a cylinder head. Pistons are mounted in the cylinder block so as to be movable back and forth axially. The compressor drive is provided via a belt pulley by means of a drive shaft 1. The compressors in the present case are compressors having variable piston stroke, the piston stroke being regulated by a pressure difference defined by the pressures on a gas inlet side and in a drive mechanism chamber. Depending on the magnitude of the pressure difference, a tilt plate in the form of a tilt ring 2 is deflected, or tilted, from its vertical position to a greater or lesser degree. The greater the resulting angle of tilt, the greater is the piston stroke and, therefore, the higher is the pressure made available on the outlet side of the compressor.

From FIG. 1 it can be seen that the tilt plate mechanism of a first preferred embodiment of a compressor according to the invention comprises the tilt ring 2, a sliding sleeve 3, which is mounted so as to be axially displaceable on the drive shaft, a spring 4, a gas force support means 5 composed of a supporting element 6 and a force transmission element 7, a securing element 8 and drive pins 9 which serve for transmitting torque between the drive shaft 1 and the tilt ring 2.

The supporting element in the present first preferred embodiment is of cylindrical or barrel-like shape. The force transmission element is mounted, so as to be rotatable about its longitudinal axis, in a corresponding recess 10 in the drive shaft. As already indicated by the terminology, the gas force support means 6 serves substantially only for providing axial support for the piston forces, whereas the transmission of torque to the tilt plate is accomplished substantially by the drive pins 9.

The sliding sleeve 3 has two flattened sides 11 (only one flattened side can be seen in FIG. 1), which are in sliding engagement with corresponding flattened regions 12 on the tilt ring 2. The force transmission element 7 has a shoulder 13, which governs its position (especially in the radial direction) in the drive shaft 1. On that side of the force transmission element 7 which is remote from the supporting element 6, the securing element 8 ensures that the gas force support means 5, or that is to say the supporting element 6 and the force transmission element 7, remains securely in the drive shaft. In addition to the connection between the tilt ring 2 and the drive shaft 1 already explained hereinbefore, the drive pins 9 also provide the connection between the sliding sleeve 3 and the drive shaft 1 and resultant force/torque transmission. The drive pins 9 project into a recess in the drive in the drive shaft in the form of grooves 14 (again, only one of the grooves 14 can be seen in FIG. 1). The drive pins are introduced into corresponding recesses 16 in the tilt ring 2 with a press fit.

The spring 4 serves as a connection element, arranged between the drive shaft 1 and the sliding sleeve 3, allowing forces to be transmitted in the axial direction. It is mounted so as to be axially displaceable on the drive shaft 1. That end of the force transmission element 7 which is remote from the supporting element 6 projects through a longitudinal slot 17 formed in the sliding sleeve 3 and into the drive shaft 1. At this juncture it should be pointed out that, as an alternative to or also in addition to the force/torque transmission by way of the drive pins 9, the sliding sleeve can be so constructed that a longitudinal slot arranged opposite the longitudinal slot 17 is provided in the sliding sleeve, into which slot that end of the force transmission element 7 which is remote from the supporting element 6 projects, consequently transferring drive torque from the drive shaft 1 to the sliding sleeve 3. It should again be briefly mentioned at this juncture that the drive shaft 1 and the sliding sleeve 3 can, in addition to or as an alternative to the connection and/or torque transmission by way of the drive pins 9, have flattened regions that correspond to one another so that the sliding sleeve is mounted on the drive shaft for conjoint rotation therewith (not shown in FIG. 1).

The arrangement shown in an exploded view in FIG. 1 is shown again in longitudinal section in FIG. 2, it being the case that, in addition to the features already known from FIG. 1, it can be seen from FIG. 2 how the tilt ring 2 is mounted in a holder 18 connected to piston rods or the pistons. Sliding blocks 19 located between the tilt ring 2 and the holder 18 serve for mounting. In the view shown in FIG. 2, the tilt ring 2 is in a position of minimum deflection, that is to say the tilt angle of the tilt ring is at a minimum. This view shows especially the co-operation of the shoulder 13 formed on the gas force support means 5 with the drive shaft 1. There can furthermore also be seen the interplay between the force transmission element 7 and the securing element 8.

FIG. 3 a shows a view corresponding to FIG. 2, that is to say a longitudinal section through the first preferred embodiment, but in this case at a maximum deflection angle of the tilt ring 2. FIG. 3 b furthermore shows a transverse section through the tilt ring mechanism according to FIG. 3 a. It should again be pointed out at this juncture that, in the preferred embodiment, the drive torque is provided by the drive pins 9 (indicated by the arrow 25) and not by the gas force support means 5 and supporting element 6 (indicated by the arrows 23, 24). The drive torque is indicated in FIG. 3 b by an arrow 20. Furthermore, the tilt axis 21 of the tilt ring 2 can be seen in FIG. 3 b.

FIG. 4 a again provides a detailed view of parts of the force transmission element 7 and the supporting element 6 in engagement with the tilt ring 2. FIG. 4 b shows two longitudinal sections through the gas force support means 5 of the first preferred embodiment which are produced by rotation through 90° from one another. This view clearly shows that the supporting element 6 is of cylindrical shape. Because the contour of cylindrical or barrel-like shape extends perpendicular to the tilt plane to a degree which is not negligible, a torsional moment (which acts perpendicular to the tilt moment of the tilt ring and which is brought about inter alia because the maximum gas force at a piston occurs at the moment of opening of the valve and not in the dead-centre of the piston) can be introduced there, that is to say at the cylindrical supporting element 6, unless the latter is mounted in the drive shaft 1 so as to be rotatable about its central axis in a manner in accordance with the invention. For that reason, an arrangement according to the invention ensures that the torsional moment is introduced only into the elements provided for the purpose, which can be, for example, the spindle-like drive pins 9 or else any desired supporting surfaces. Introduction of the torsional moment into the gas force support means 5 is ruled out by an arrangement according to the invention.

At this juncture there should briefly be mentioned again the advantages of the invention, which are as follows: the gas force support means 5, in substantially and preferably torque-free manner (provided that an arrangement is selected in which that end of the force transmission element 7 which is remote from the supporting element 6 is not in torque-transferring engagement with the sliding sleeve 3 at its end which is remote from the supporting element 6), takes on the supporting function of the tilt ring 2 in respect of the axially acting piston forces; the supporting element 6 or at least the head region of the supporting element 6 can be formed so as to have a large area. that is to say a cylindrical or barrel shape, it not being possible for torsional moments to be introduced because the gas force support means 6 can orientate itself about its central axis; the drive moments are transmitted in defined manner in the plane perpendicular to the tilting plane of the tilt ring, in which context it should be pointed out that there are various possibilities for force and torque transmission. At this juncture reference should again be made to FIGS. 4 a and 4 b, in which there is shown a gas force support means 5 which has a cylindrical supporting element 6. The supporting element 6 is in engagement with the tilt ring 2. As a result of the fact that the force transmission element 7 of the gas force support means 5 is mounted in the drive shaft 1 so that it can rotate about its own axis, substantially no torsional moment (twisting) can be transmitted. This allows defined transmission of the torsional moment at another location, as already mentioned hereinbefore, and prevents restriction of the mechanism. This also allows simple and rapid assembly. Over-determination in respect of the torsional moment, which could result in the case of the proposed cylindrical formation of the supporting element 6 of the gas force support means 5, is avoided as a result of the latter's being mounted in the drive shaft 1 so that it is rotatable about its own axis. The forces in the direction of the recess 10 in the drive shaft 1 are transmitted by the shoulder 13 of the gas force support means 5 and by the securing element 8 at the other end of the gas force support means 5, or that is to say at that end of the force transmission element 7 which is remote from the supporting element 6. From FIG. 3 a and, especially, FIG. 3 b it can be seen that the opening in the tilt ring 2 in which the supporting element 6 engages is so formed that the supporting element 6 cannot transmit any driving torque. As a result, as already mentioned hereinbefore, the loading on the gas force support means 5 is less and the drive torque can be transmitted in defined manner and at another location (in the present embodiment, at the drive pins 9).

Hereinbelow there will be given further details of the transmission of the drive torque: As already mentioned in the description of FIG. 1, the tilt ring 2 is connected to the sliding sleeve 3 and to the drive shaft 1 by means of the drive pins 9. The sliding sleeve 3 is mounted so as to be axially displaceable on the drive shaft 1 and, in co-operation with the spring 4, the drive pins 9 and the gas force support means 5, makes possible the adjustment of the tilt angle of the tilt ring 2. The tilt angle brought about depends on the gas forces, the inertia characteristics of the tilt ring 2 and the pistons in engagement therewith, and also on the spring force of the spring 4. The sum of the moments about the tilt axis is, in other words, equal to zero (tilt moments equal to zero). The drive pins 9 are secured axially against dropping out, which is accomplished by means of a press fit of the pins in the sliding sleeve 3 or tilt ring 2. In the present preferred embodiment, the transmission of the drive torque from the drive shaft 1 to the tilt ring 2 is accomplished directly by means of the drive pins 9. Alternatively, it is feasible for the drive torque to be transmitted indirectly by way of the sliding sleeve 3. In both cases, however, there are elements (for example, drive pins 9) which are connected to, or which project into, the shaft. Of course it is also feasible for there to be only one element. As a result, the radial orientation of the sliding sleeve 3 is fixed and, by means of a sufficiently large recess in the sliding sleeve, it is ensured that that part of the gas force support means 5 or force transmission element 7 which faces the supporting element 6 cannot transmit any moment to the sliding sleeve. In FIGS. 3 a and 3 b there is shown an example in which the drive pins 9, which are connected to the tilt ring 2, project into a groove 14 in the drive shaft 1. As a result, the drive torque is transmitted directly by the drive pins 9 from the drive shaft 1 to the tilt ring 2.

Alternatively, indirect transmission of the drive torque with a force transfer by way of the sliding sleeve is feasible. This could be accomplished in constructional terms as follows: a connecting element—between the drive shaft 1 and the sliding sleeve 3—which allows the transmission of forces and moments in the radial direction but which allows the axial displaceability of the bush, for example by sliding in a groove in the sliding sleeve 3. Such a connecting element could be, for example, a feather key. That end of the force transmission element 7 which is remote from the supporting element 6 is passed through the shaft and projects into a slot in the sliding sleeve 3, in which slot the force transmission element 7 is tightly guided and as a result can transmit the drive torque.

A central point of the present invention is the formation of the gas force support means 5. In the context of the present invention, a gas force support means is provided which, on the one hand, is relieved of loading as a result of its not transmitting any drive torque but which, on the other hand, is optimised in respect of the contact pressure due to transmission of the gas forces.

The recess 22 in the tilt ring 2, in which recess the supporting element 6 engages, is so formed that the gas force support means 5 and especially the supporting element 6 are radially free and accordingly do not transmit any drive torque. Furthermore, the gas force support means 5 and the supporting element 6 are so formed that tilting of the tilt ring 2 is accomplished by means of a process of rolling on the supporting element 6. Ideally, the height of the recess 22 for the supporting element 6 does not change. This simplifies machining of the recess 22, although a change in the height of the recess 22 is theoretically also possible.

Taking into account the aspects explained hereinbefore in greater detail, the supporting element 6 is optimally constructed when it allows the rolling movement by means of a suitable curve profile and continues linear in the radial direction so that line contact and low Hertzian stress are ensured.

An example of a corresponding gas force support means 5, especially for a corresponding supporting element 6, is shown in FIGS. 4 a and 4 b. In this example, the curved shape is a circle so that the supporting element is of cylindrical shape. In addition, the centre-point of the circle, or that is to say of the circular shape of the cylinder, coincides with the centre-point of the piston articulation, which results in a constant clearance volume of the compressor according to the invention. The following relations further apply to FIG. 4 a: The wall thicknesses of the tilt ring, or that is to say of the walls S2 and S3 surrounding the recess 22, are equal (S2=S3) and the radii of curvature R1 and R2 of a radially outer (R1) and radially inner (R2) contour are, by virtue of the cylindrical shape of the supporting element 6, also equal (R1=R2). It furthermore holds true that the height S1, or that is to say width, of the recess 22 is equal to the sum of the two radii of curvature R1 and R2 (S1=R1+R2). Furthermore, this sum is constant. The centre-points of R1, R2 and R, wherein R denotes the radius of the sliding blocks 19, are identical, which ensures a constant clearance volume.

In principle, however, the curve profile of the gas force support means can be freely selected subject to the following preconditions: For each tilt angle, the height S1 of the recess must be equal to the sum of the radii of curvature of the radially outer R1 and the radially inner R2 contour. Simple examples of profiles which roll in a recess 22 of the same height S1 are shown in FIGS. 5 and 6.

FIG. 5 (the second preferred embodiment of a compressor according to the invention) shows how, by selecting different radii whilst maintaining the clearance volume constant, it is possible to increase the wall thickness of the tilt ring 2 in the region of the recess 22 on that side which is subject to major loading by the gas force (the side facing the pistons). For the arrangement according to FIG. 5, the following relations apply: S2>S3, R1<R2, S1=R1+R2=constant (in the region of tilting), which results in a constant clearance volume.

A third preferred embodiment is depicted in FIG. 6, which shows an example in which the remaining wall thicknesses of the tilt ring 2 in the region of the recess 22 are of equal size (S2=S3). The main loading resulting from the compression force acts in this example on the line contact between the tilt ring 2 and a surface which is cylindrical in parts and is formed having the radius R1. The condition R1>R2 results in the advantage of less contact pressure for the main loading direction. At this juncture, however, it should be pointed out that, in this embodiment, the central point of the articulation (centre-point of radii R1 and R2) in the tilt ring 2 does not coincide with the central point of the piston articulation, which is indicated by a spacing S4 not being equal to zero. This means that the clearance volume is not constant. In summary, the following relations apply to FIG. 6: S3=S2, R1>R2 (which results in reduced contact pressure), S1=R1+R2=constant (in the region of tilting), the clearance volume is not constant.

In addition to designing the gas force support means 5 with regard to contact pressure, the present invention also simplifies, as already mentioned hereinbefore, assembly of the parts and their machinability. As also already mentioned hereinbefore, the gas force support means 5 and especially the force transmission element 7 are not fixedly connected to the drive shaft 1 but rather are mounted in rotatable manner therein. The orientation of the tilt axis is predetermined by the drive pins 9 and the flattened regions 11 on the sliding sleeve 3 and the flattened regions 12 on the tilt ring (cf. FIGS. 7 a and 7 b in this regard). A gas force support means 5 having a fixed press fit in the drive shaft 1 would also predetermine the orientation of the tilt axis and accordingly create an over-determination which would, at the least, cause difficulties for assembly and problem-free tilting of the mechanism. Such an arrangement is possible only with extremely accurate adjustment of play and extremely precise assembly and would considerably increase machining and assembly costs. The present invention avoids this over-determination as a result of the rotatable mounting of the force transmission element 7 and gas force support means 5 in the drive shaft 1 and facilitates assembly.

Finally, there should be mentioned, as a substantial advantage of the rotatability of the gas force support means 5, the circumstance that the moments acting with respect to the axis of rotation cannot be taken up. From this it follows that the forces resulting from the torsional moment are substantially not taken up by the gas force support means 5 (cf. FIGS. 7 a and 7 b in this respect). The flattened regions 12 of the tilt ring 2 and the flattened regions 11 of the sliding sleeve 3 are, by virtue of the lever arm conditions, substantially better suited to taking up the forces resulting from the torsional moment. As FIGS. 7 a and b show, the force is transferred from the flattened region of the tilt ring via the sliding sleeve 3 to the drive shaft 1. Furthermore, the axis 26 for the torsional moment and the application point 27 for the resulting compression force are indicated in FIGS. 7 a and b.

Although the invention is described using embodiments having fixed combinations of features, it nevertheless also encompasses any further feasible advantageous combinations of those features, as are especially but not exhaustively mentioned in the subordinate claims. All features disclosed in the application documents are claimed as being important to the invention insofar as they are novel on their own or in combination compared with the prior art.

REFERENCE NUMERALS

-   1 drive shaft -   2 tilt ring -   3 sliding sleeve -   4 spring -   5 gas force support means -   6 supporting element -   7 force transmission element -   8 securing element -   9 drive pin -   10 recess in drive shaft 1 -   11 flattened side of sliding sleeve 3 -   12 flattened region on tilt ring 2 -   13 shoulder -   14 groove -   15 recess in sliding sleeve 3 -   16 recess in tilt ring 2 -   17 longitudinal slot -   18 holder -   19 sliding block -   20 arrow -   21 tilt axis -   22 recess in tilt ring 2 for accommodating supporting element 6 -   23 arrow -   24 arrow -   25 arrow -   26 axis -   27 application point 

1. Axial piston compressor, especially for motor vehicle air-conditioning systems, having a tilt plate (2), especially a ring-shaped tilt plate, which is variable in terms of its inclination with respect to a drive shaft (1) and which is driven in rotation by the drive shaft (1) and is in articulated connection with at least one supporting element (6) arranged at a spacing from the drive shaft (1) and rotating together therewith, the pistons in each case having an articulated arrangement with which the tilt plate (2) is in sliding engagement, and the supporting element (6) being arranged at the radially outer end of a force transmission element (7) which rotates together with the drive shaft (1) and is fixed in the latter so as to be non-displaceable in the radial direction, characterised in that the force transmission element (7) is mounted in the drive shaft (1) so that it can rotate about its longitudinal axis.
 2. Compressor according to claim 1, characterised in that the supporting element (6) and force transmission element (7) serve substantially only for providing axial support for the pistons or support for the gas force, whereas an arrangement independent thereof, especially an articulated connection between the drive shaft (1) and the tilt plate (2), serves substantially only for torque transmission.
 3. Compressor according to claim 1, characterised in that the supporting element (6) has a basic shape which in radial section is approximately rectangular, the corners being highly rounded especially with different radii, or compressed or circular or in the shape of a deformed circle or ellipsoidal.
 4. Compressor according to claim 3, characterised in that those regions of the supporting element (6) which are in contact with the tilt plate (2) are of at least partly cylindrical or barrel-shaped construction.
 5. Compressor according to claim 1, wherein the tilt plate (2) is pivotally mounted on a sliding sleeve (3) mounted so as to be axially displaceable along the drive shaft (1), characterised in that between the sliding sleeve (3) and the tilt plate (2) there is provided an arrangement, especially at least one cylindrical-pin-like element or supporting or contact surfaces, in order to provide support for a torsional moment acting in the region of the drive shaft (1).
 6. Compressor according to claim 1, characterised in that the force transmission element (7) comprises, at least over parts of its periphery, a shoulder (13) in the region of the drive shaft (1) and/or, at its end remote from the supporting element (6), a securing element (8), especially extending in the axial direction.
 7. Compressor according to claim 5, characterised in that the tilt plate (2) is connected by way of drive pins (9) to the sliding sleeve (3) and/or to the drive shaft (1).
 8. Compressor according to claim 7, characterised in that the drive pins (9) are introduced into the sliding sleeve (3) or the tilt plate (2) with a press fit.
 9. Compressor according to claim 8, characterised in that the drive pins (9) project into a recess (22), especially a groove (14), in the drive shaft (1).
 10. Compressor according to claim 5, characterised in that a connecting element, especially a feather key, is arranged between the drive shaft (1) and the sliding sleeve (3), which connecting element allows transmission of forces and moments in a radial direction and is mounted in axially displaceable manner on the drive shaft (1).
 11. Compressor according to claim 10, characterised in that that end of the force transmission element (7) which is remote from the supporting element (6) projects through the drive shaft (1) and into a longitudinal slot (17) in the sliding sleeve (3) in such a way that drive torque is transmitted from the drive shaft (1) to the sliding sleeve (3) by means of that end of the force transmission element (7) which is remote from the supporting element (6).
 12. Compressor according to claim 1, characterised in that the supporting element (6) is so constructed that it is located within a recess (22) in the tilt plate (2) in line contact with the latter.
 13. Compressor according to claim 12, characterised in that the height (S1) of a/the recess (22) in the tilt plate (2) is equal to the sum of the radii of curvature of a radially outer (R1) and radially inner (R2) contour.
 14. Compressor according to claim 13, characterised in that, on that side of the tilt plate (2) which faces the pistons, the wall thickness (S2) in the region of the/a recess (22) in the tilt plate (2) is greater than on that side which is remote from the pistons (S2>S3), whilst at the same time the clearance volume is constant for all tilt angles of the tilt plate (2).
 15. Compressor according to claim 10, characterised in that the drive shaft (1) and the sliding sleeve (3) have flattened regions (11) that correspond to one another, so that the sliding sleeve (3) is mounted on the drive shaft (1) in a manner ensuring conjoint rotation.
 16. Compressor according to claim 15, characterised in that the tilt plate (2) has at least one flattened region (12) which corresponds to a flattened region (11) on the sliding sleeve (3). 